Axial turbine wheel

ABSTRACT

A turbocharger including a turbine wheel having a hub-to-tip ratio of no more than 60% and blades with a high turning angle, a turbine housing forming an inwardly spiraling primary-scroll passageway that significantly converges to produce highly accelerated airflow into the turbine at high circumferential angles, and a two-sided parallel compressor. The compressor and turbine each produce substantially no axial force, allowing the use of minimal axial thrust bearings.

The present invention relates generally to turbochargers and, moreparticularly, to an axial turbine having a low hub to tip ratio.

BACKGROUND OF THE INVENTION

With reference to FIG. 1, a typical turbocharger 101 having a radialturbine includes a turbocharger housing and a rotor configured to rotatewithin the turbocharger housing along an axis of rotor rotation 103 onthrust bearings and two sets of journal bearings (one for eachrespective rotor wheel), or alternatively, other similarly supportivebearings. The turbocharger housing includes a turbine housing 105, acompressor housing 107, and a bearing housing 109 (i.e., a centerhousing that contains the bearings) that connects the turbine housing tothe compressor housing. The rotor includes a turbine wheel 111 locatedsubstantially within the turbine housing, a compressor wheel 113 locatedsubstantially within the compressor housing, and a shaft 115 extendingalong the axis of rotor rotation, through the bearing housing, toconnect the turbine wheel to the compressor wheel.

The turbine housing 105 and turbine wheel 111 form a turbine configuredto circumferentially receive a high-pressure and high-temperatureexhaust gas stream 121 from an engine, e.g., from an exhaust manifold123 of an internal combustion engine 125. The turbine wheel (and thusthe rotor) is driven in rotation around the axis of rotor rotation 103by the high-pressure and high-temperature exhaust gas stream, whichbecomes a lower-pressure and lower-temperature exhaust gas stream 127and is axially released into an exhaust system (not shown).

The compressor housing 107 and compressor wheel 113 form a compressorstage. The compressor wheel, being driven in rotation by the exhaust-gasdriven turbine wheel 111, is configured to compress axially receivedinput air (e.g., ambient air 131, or already-pressurized air from aprevious-stage in a multi-stage compressor) into a pressurized airstream 133 that is ejected circumferentially from the compressor. Due tothe compression process, the pressurized air stream is characterized byan increased temperature over that of the input air.

Optionally, the pressurized air stream may be channeled through aconvectively cooled charge air cooler 135 configured to dissipate heatfrom the pressurized air stream, increasing its density. The resultingcooled and pressurized output air stream 137 is channeled into an intakemanifold 139 on the internal combustion engine, or alternatively, into asubsequent-stage, in-series compressor. The operation of the system iscontrolled by an ECU 151 (engine control unit) that connects to theremainder of the system via communication connections 153.

U.S. Pat. No. 4,850,820, dated Jul. 25, 1989, which is incorporatedherein by reference for all purposes, discloses a turbocharger similarto that of FIG. 1, but which has an axial turbine. The axial turbineinherently has a lower moment of inertia, reducing the amount of energyrequired to accelerate the turbine. As can be seen in FIG. 2, theturbine has a scroll that circumferentially receives exhaust gas at theradius of the turbine blades and (with reference to FIG. 1) axiallyrestricts the flow to transition it to axial flow. It thus impacts theleading edge of the turbine blades in a generally axial direction (withreference to col. 2).

For many turbine sizes of interest, axial turbines typically operate athigher mass flows and lower expansion ratios than comparable radialturbines. While conventional axial turbines generally offer a lowerinertia, albeit with some loss of efficiency and performance, theysuffer from an inability to be efficiently manufactured in the smallsizes usable with many modern internal combustion engines. This is,e.g., due to the exceptionally tight tolerances that would be required,due to aerodynamic limitations, and/or due to dimensional limitations oncreating small cast parts. Axial turbines also lack the ability toperform well at higher expansion ratios, such as are typically neededdue to the pulsing nature of the exhaust of an internal combustionengine. Furthermore, conventional axial turbines have a significantchange in static pressure across the blades, causing significant thrustloads on the thrust bearings of the rotor, and potentially causingblowby.

In some conventional turbochargers the turbines and compressors areconfigured to exert axial loads in opposite directions so as to lessenthe average axial loads that must be carried by the bearings.Nevertheless, the axial loads from the turbines and compressors do notvary evenly with one another and may be at significantly differentlevels, so the thrust bearings must be designed for the largest loadcondition that may occur during turbocharger use. Bearings configured tosupport high axial loads waste more energy than comparable low-loadbearings, and thus turbochargers that must support higher axial loadslose more energy to their bearings.

Accordingly, there has existed a need for a turbocharger turbine havinga low moment of inertia, and characterized by a small size that does notrequire exceptionally tight tolerances, while having reasonableefficiency both at both lower and higher expansion ratios, and smalleraxial loads. Preferred embodiments of the present invention satisfythese and other needs, and provide further related advantages.

SUMMARY OF THE INVENTION

In various embodiments, the present invention solves some or all of theneeds mentioned above, typically providing a cost effective turbochargerturbine characterized by a low moment of inertia, and having a smallsize that does not require exceptionally tight tolerances, whileoperating at reasonable efficiency levels at both at both lower andhigher expansion ratios, and having only small changes in static loads.

The invention provides a turbocharger configured to receive an exhaustgas stream from an engine configured to operate over a range of standardoperating conditions, and to compress input air into a pressurized airstream. The turbocharger includes a turbocharger housing including aturbine housing, and a rotor configured to rotate within theturbocharger housing along an axis of rotor rotation. The rotor includesan axial turbine wheel, a compressor wheel, and a shaft extending alongthe axis of rotor rotation and connecting the turbine wheel to thecompressor wheel. The turbine wheel is configured with a hub and aplurality of axial turbine blades configured to drive the rotor inrotation around the axis of rotor rotation when the turbochargerreceives exhaust gas stream from the engine from a circumferentialdirection. The compressor wheel is configured to compress input air intothe pressurized air stream.

Advantageously, the turbine housing forms an inwardly spiraling turbineprimary-scroll passageway characterized by a significant enough radialreduction to accelerate exhaust gas such that a significant portion ofthe total pressure of the exhaust gas received by the turbine isconverted into dynamic pressure. This allows an appropriately configuredblade to extract a significant amount of energy from the exhaust gaswithout significantly changing the static pressure across the turbineblades. With a substantially unchanged static pressure across theturbine blades, the exhaust gas stream applies little to no axialpressure on the rotor.

The turbine wheel blades have an axially upstream edge, an axiallydownstream edge, a hub end, and a tip end opposite the hub end. Thetrailing edge is characterized by a radius at the hub end and a radiusat the tip end. A feature of the invention is that the radius at the hubend of the turbine wheel trailing edge is no more than 60% of the radiusof the tip end of the turbine wheel trailing edge. Further featuresinclude that the turbine wheel blades are limited to 16 or less innumber, and are each characterized by a large turning angle.

Advantageously, these features provide for the extraction of asignificant amount of energy from high-speed exhaust gas received in ahighly circumferential direction without significantly impacting thestatic pressure of the gas. Furthermore, the turbine wheel does notrequire extremely tight manufacturing tolerances or small blade sizes,even when the wheel is manufactured in relatively small sizes.

The invention further features that the compressor may be a two-sided,parallel, radial compressor including a compressor wheel withback-to-back oriented impeller blades including a first set of impellerblades facing axially away from the turbine and a second set of impellerblades facing axially toward the turbine. The compressor housing isconfigured to direct inlet air to each set of compressor blades inparallel. Advantageously, under this feature, the compressor isconfigured to produce substantially no axial load on the rotor. Incombination with a turbine that also produces little or no axial load onthe rotor, thrust bearing load levels can be significantly lower than inconventional turbochargers. The lower bearing load levels allow for theuse of a more efficient thrust bearing, and thus increase the resultingoverall efficiency of the turbocharger.

Other features and advantages of the invention will become apparent fromthe following detailed description of the preferred embodiments, takenwith the accompanying drawings, which illustrate, by way of example, theprinciples of the invention. The detailed description of particularpreferred embodiments, as set out below to enable one to build and usean embodiment of the invention, are not intended to limit the enumeratedclaims, but rather, they are intended to serve as particular examples ofthe claimed invention.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a system view of a prior art turbocharged internal combustionengine.

FIG. 2 is a cross-sectional plan view of a turbocharger embodying thepresent invention.

FIG. 3 is a cross-sectional side view of the turbocharger depicted inFIG. 2, taken along line A-A of FIG. 2.

FIG. 4 is a plan view of certain critical flow locations relative to aturbine wheel depicted in FIG. 2.

FIG. 5 is a depiction of the camber of a turbine blade depicted in FIG.2.

FIG. 6 is a perspective view of the turbine wheel depicted in FIG. 2.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

The invention summarized above and defined by the enumerated claims maybe better understood by referring to the following detailed description,which should be read with the accompanying drawings. This detaileddescription of particular preferred embodiments of the invention, setout below to enable one to build and use particular implementations ofthe invention, is not intended to limit the enumerated claims, butrather, it is intended to provide particular examples of them.

Typical embodiments of the present invention reside in a motor vehicleequipped with a gasoline powered internal combustion engine (“ICE”) anda turbocharger. The turbocharger is equipped with a unique combinationof features that may, in various embodiments, provide the aerodynamicbenefits of a zero reaction turbine with the geometric benefits of afifty percent reaction turbine, and/or provide significantly improvedsystem efficiencies by combining less efficient components in a mannerthat reduces the bearing requirements, and thereby forms a system with ahigher efficiency than the comparable unimproved system.

The turbine is configured to operate at reasonable efficiency levels atboth lower and higher expansion ratios, having only small changes instatic pressure across the turbine wheel (and thereby low rotor thrustloads), while it has a low moment of inertia, and is characterized by asmall size, but does not require exceptionally tight tolerances. Incombination with this, the compressor is also characterized by low axialthrust loads, providing for the turbocharger to require a thrust bearingthat is significantly more efficient than is used in comparableconventional turbochargers.

With reference to FIGS. 2 & 3, in a first embodiment of the invention atypical internal combustion engine and ECU (and optionally anintercooler), such as are depicted in FIG. 1, are provided with aturbocharger 201 that includes a turbocharger housing and a rotorconfigured to rotate within the turbocharger housing along an axis ofrotor rotation 203 on a set of bearings. The turbocharger housingincludes a turbine housing 205, a compressor housing 207, and a bearinghousing 209 (i.e., a center housing that contains radial and thrustbearings) that connects the turbine housing to the compressor housing.The rotor includes an axial turbine wheel 211 located substantiallywithin the turbine housing, a radial compressor wheel 213 locatedsubstantially within the compressor housing, and a shaft 215 extendingalong the axis of rotor rotation, through the bearing housing, toconnect the turbine wheel to the compressor wheel and provide for theturbine wheel to drive the compressor wheel in rotation around the axisof rotation.

The turbine housing 205 and turbine wheel 211 form a turbine configuredto circumferentially receive a high-pressure and high-temperatureexhaust gas stream from an exhaust manifold of the engine (such as theexhaust gas stream 121 from the exhaust gas manifold 123, as depicted inFIG. 1). The turbine wheel (and thus the rotor) is driven in rotationaround the axis of rotor rotation 203 by the high-pressure andhigh-temperature exhaust gas stream acting on a plurality of blades 231of the turbine wheel. The exhaust gas stream becomes a lower totalpressure exhaust gas stream while passing through the blades, and issubsequently axially released via a turbine outlet 227 into an exhaustsystem (not shown).

The compressor housing 207 and compressor wheel 213 form a radialcompressor. The compressor wheel, being driven in rotation by theexhaust-gas driven turbine wheel 211 (via the shaft 215), is configuredto compress axially received input air (e.g., ambient air, oralready-pressurized air from a previous-stage in a multi-stagecompressor) into a pressurized air stream that may be ejectedcircumferentially from the compressor and sent on to an engine inlet(such as pressurized air stream 133 that is sent on to the engine inlet139, as depicted in FIG. 1).

Turbine Volute

The turbine housing 205 forms an exhaust gas entrance passageway 217leading into a primary-scroll passageway 219 configured to receive theexhaust gas stream from the engine in a direction normal to and radiallyoffset from the rotor axis of rotation 203. The primary-scrollpassageway forms a spiral adapted to significantly accelerate the speedof the gas stream to a high speed, which may be a supersonic speed forat least some operating conditions of the turbine (and its relatedengine). More particularly, the primary-scroll passageway turns theexhaust gas both inwardly around the axis of rotation 203 and axiallytoward the axial turbine wheel 211, thereby achieving (for some standardoperating conditions of the engine) a supersonic flow having both adownstream axial component 221 and a downstream circumferentialcomponent 223.

Effectively, this configuration takes advantage of the conservation ofangular momentum (rather than a convergent divergent nozzle) to achievea high-speed airflow that may include a shockless transition tosupersonic speeds for at least some operating conditions. Typically, aspiral characterized by a large radius change is required to achievethis change in velocity, and even though the resulting airstream isturned axially into an axial turbine wheel, it has a very high-speedcircumferential component.

This circumferential component is achieved without the use of turningvanes, which would cause additional losses. Thus, the turbine inlet ofthis embodiment is of a vaneless design. As compared to a design withvanes, such a design advantageously is cost efficient, reliable (in thatit eliminates parts from an environment in which they are likely toerode), avoids friction pressure losses, and avoids establishing acritical throat area that could choke the flow in some operatingconditions.

With reference to FIGS. 2-4, this potentially supersonic flow of theaccelerated exhaust gas stream in the inner radius of the primary-scrollpassageway is directed into the turbine wheel 211. More particularly,the primary-scroll passageway is an inwardly spiraling passagewaycharacterized by a primary-scroll inlet port 225 that connects theprimary-scroll passageway to the exhaust gas entrance passageway 217.The primary-scroll passageway substantially forms a convergentpassageway that spirals inward enough and converges enough to acceleratethe exhaust gas, and to achieve supersonic speeds for at least somestandard operating conditions of the engine (and thus of theturbocharger) as the exhaust gas turns axially downstream and impingeson the axially upstream end 233 of the blades 231.

The primary-scroll inlet port 225 is a planar location located along thepassageways within the turbine that the exhaust gas travels throughprior to reaching the turbine wheel. The location of the primary-scrollinlet port is defined relative to an opening in the passageway, which ischaracterized by a tongue-like shape when viewed in a cross-sectiontaken normal to the rotor axis of rotation 203.

More particularly, the structure of a tongue 235 appears as a protrusionhaving a tip when viewed in the cross-section of FIG. 3. It should benoted that in some embodiments this structure will not vary in shapewhen the cross section is taken at different axial locations. In otherembodiments the structure forming the tongue 235 may be shaped such thatthe location of the tip of the tongue varies when viewed incross-sections taken at different axial locations.

The primary-scroll inlet port 225 is located at the tip of the tongue235. To any extent that the circumferential location of the tip of thetongue appears to vary with the axial location of the cross-sectionconsidered, the primary-scroll inlet port 225 is defined to be at themost upstream location of the tip of the tongue, i.e., the upstream-mostlocation at which the housing opens such that it is no longer radiallyinterposed between the exhaust gas stream and the blades (even thoughthe blades are axially offset from the exhaust gas stream). For thepurposes of this application, the primary-scroll inlet port 225 isdefined as the smallest planar opening from the exhaust gas entrancepassageway 217 into the primary-scroll passageway 219, at the tip of thetongue. In other words, it is at the downstream end of the exhaust gasentrance passageway at the location at which the stream opens up to theblades.

The primary-scroll passageway 219 starts at the primary-scroll inletport 225, and spirals inward 360 degrees around the axis of rotation toform a converging loop that rejoins flow coming in the primary-scrollinlet port 225. This convergent loop accelerates the exhaust gascircumferentially and turns it axially. Throughout the 360 degrees ofthe primary-scroll passageway 219, the accelerated and turned exhaustgas stream impinges on the blades 231, passing between the blades anddriving the turbine wheel 211 in rotation.

In summary, the housing for the axial turbine wheel forms an inwardlyspiraling primary-scroll passageway that surrounds the axis of rotorrotation. It begins at a primary-scroll inlet port 225 that issubstantially radially external to the axially upstream ends of theblades, providing for the passageway to spiral inwardly and turn axiallyto accelerate the exhaust gas flow into the upstream ends of the axialturbine wheel blades.

Corrected Mass Flow

To provide for an adequate level of acceleration of the exhaust gasunder the invention, the primary-scroll passageway 219 is configuredwith sizing parameters such that the corrected mass flow rate surfacedensity of the turbine, when operated at a critical expansion ratio(E_(cr)), exceeds a critical configuration parameter, i.e., a criticalcorrected mass flow rate surface density (D_(cr)). More particularly,the sizing parameters for the scroll include a primary-scroll radiusratio (r_(r)) and a primary-scroll inlet port area (a_(i)), and areselected such that the corrected mass flow rate surface density of theturbine exceeds the critical configuration parameter D_(cr) when theturbine is operated at the critical expansion ratio E_(cr). These sizingparameters are defined relative to the primary-scroll inlet port 225,which is characterized by a centroid 237. For the gas to be axiallyadequately accelerated, this centroid will be substantially radiallyexternal to, and typically axially upstream of, an axially upstream end233 of each blade 231.

The values of some of the above-recited terms are dependent upon thetype of exhaust stream gas that will be driving the turbine. Thisexhaust-stream gas will be characterized by a Boltzmann Constant (k),and by a Gas Constant R-specific (R_(sp)). These constants vary by gastype, but for most gasoline powered engine exhaust gasses, thedifference is anticipated to be small, with the constants beingtypically be on the order of k=1.3 and R_(sp)=290.8 J/kg/K.

The turbine housing has an ability to accelerate the exhaust gas that ischaracterized by the two sizing parameters recited above. The firstsizing parameter, being the primary-scroll radius ratio r_(r) is definedto be a radius of a point 239 at the hub at leading edge of the turbineblades 231 (i.e., at the inner edge of the rotor inlet), divided by aradius of the centroid 237 of the planar area of the primary-scrollinlet port 225. The second, being the primary-scroll inlet port areaa_(i) is defined to be the area of the primary-scroll inlet port 225.

As mentioned above, the geometry of this embodiment of a turbine isdefined relative to operational parameters at the critical expansionratio E_(cr). This critical expansion ratio is obtained from the formula

$E_{cr} = \left( \frac{k + 1}{2} \right)^{(\frac{k}{k - 1})}$

and is a function of the gas-specific Boltzmann's Constant k. A typicalvalue for E_(cr) is 1.832.

As recited above, the dimensions of the primary-scroll passageway 219 ofthis embodiment are limited by a primary-scroll radius ratio r_(r) and aprimary-scroll inlet port area a_(i) that cause the corrected mass flowrate surface density of the turbine to exceed the critical correctedmass flow rate surface density D_(cr). This critical corrected mass flowrate surface density is obtained from the formula

$D_{cr} = {r_{r}\frac{101325}{\sqrt{288\; R_{sp}}}\left( {1 - \frac{\left( {k - 1} \right)\left( r_{r} \right)^{2}}{\left( {k + 1} \right)}} \right)^{(\frac{1}{k - 1})}\sqrt{\frac{2k}{k + 1}}}$

which varies with the primary-scroll radius ratio r_(r).

For any given turbine, exactly one steady-state inlet condition for agiven outlet static pressure (i.e., one inlet total pressure) will drivethe turbine at a given expansion ratio such as the critical expansionratio E_(cr). A variation in the geometry of the volute, e.g., avariation of the radius ratio r_(r) and/or the primary-scroll inlet portarea a_(i) can vary the steady-state mass flow rate that will drive theturbine at the given critical expansion ratio, and thus will affect therelated corrected mass flow rate surface density.

If the primary-scroll radius ratio and the primary-scroll inlet portarea are adequately selected, it will cause the corrected mass flow ratesurface density at the primary-scroll inlet port 225 when driven at thecritical expansion ratio E_(cr) to be greater than the criticalcorrected mass flow rate surface density D_(cr). While the relationshipsbetween the primary-scroll radius ratio, the primary-scroll inlet portarea and the corrected mass flow rate surface density at theprimary-scroll inlet port are complicated, and while they will typicallybe explored experimentally, it may be noted that in general, a higherradius ratio for the same port area will lead to a higher corrected massflow rate surface density.

In an iterative method of designing a turbine under the invention, aperson skilled in the art can first select a composition of an exhaustgas to be received from an engine, look up (from existing sources of gasproperties) the related Boltzmann's Constant k and Gas Constant R_(sp),and calculate the critical expansion ratio E_(cr).

A first configuration of a turbine is then designed. The turbineincludes a volute as described above, with an inwardly spiralingpassageway that turns from a tangential direction to an axial direction,and an axial turbine wheel. The design is characterized by a firstprimary-scroll radius ratio r_(r1) and a first primary-scroll inlet portarea a_(i1).

A prototype is built, put on a gas stand, and run using the selectedexhaust gas. The input total pressure is increased until a calculatedexpansion ratio reaches the critical expansion ratio E_(cr). Thisexpansion ratio is calculated from the total pressure at the inlet andthe static pressure at the outlet. A steady state mass flow rate m, atotal turbine inlet temperature T, and a total inlet pressure p_(i) aremeasured.

The corrected mass flow rate surface density is calculated from themeasured data using the following formula:

$D_{ca} = \frac{m \times \sqrt{\frac{T}{288}}}{\frac{p_{i} \times a_{i}}{101325}}$

where a_(i) is the inlet port area. This calculated corrected mass flowrate surface density D_(ca) is compared to the critical corrected massflow rate surface density D_(cr), which is calculated using thepreviously-identified formula. If the corrected mass flow rate surfacedensity exceeds or equals the critical corrected mass flow rate surfacedensity, then the design of an embodiment of the invention is complete.If the corrected mass flow rate surface density is less than thecritical corrected mass flow rate surface density, then the design isconsidered insufficient to create the high-speed circumferential airflowneeded under the invention, and another iteration of the design andtesting steps are completed.

In this next iteration, the primary-scroll radius ratio r_(r) and/or theprimary-scroll inlet port area a_(i) are appropriately adjusted (e.g.,reduced) to increase the corrected mass flow rate surface density whentaken at the critical expansion ratio E_(cr). This process is repeateduntil a design is found in which the corrected mass flow rate surfacedensity exceeds or equals the critical corrected mass flow rate surfacedensity when taken at the critical expansion ratio E_(cr).

In a potential alternative decision-making process for the above recitediterative design method, the decision to change one or both of thesizing parameters r_(r) and a_(i) is based on testing the axial loadingof the turbine wheel (or the static pressure ratios that cause axialloading) by the exhaust gasses, over critical operating conditions(i.e., conditions that cause operation to occur at the criticalexpansion ratio E_(cr)). Another iteration is conducted if the axialforces are not below a threshold, such as the loading condition when thestatic pressure upstream of the wheel near the wheel hub is greater than120% of the turbine outlet static pressure, i.e. the pressures differ atthe most by 20% of the outlet pressure.

Wheel Blades

With reference to FIGS. 3-5, relative to the downstream axial flowcomponent 221 and downstream circumferential flow component 223, eachblade 231 is characterized by a lower surface 241 (i.e., the surfacegenerally facing circumferentially into the downstream circumferentialflow component) and an upper surface 243 (i.e., the surface generallyfacing circumferentially away from the downstream circumferentialcomponent).

The lower and upper surfaces of the blade 231 meet at a leading edge 245(i.e., the upstream edge of the blade) and a trailing edge 247 (i.e.,the downstream edge of the blade). The blades extend radially outwardfrom a central hub 271 in a cantilevered configuration. They attach tothe hub along a radially inner hub end 273 of the blade, and extend to aradially outer tip end 275 of the blade. The hub end of the bladeextends from an inner, hub end of the leading edge to an inner, hub endof the trailing edge. The tip end of the blade extends from an outer,tip end of the leading edge to an outer, tip end of the trailing edge.

Typical axial turbines are typically provided with blades having bladelengths that are very small compared to the radius of the respectivehub. Contrary to this typical convention, the present embodiment isprovided with blades having a hub-to-tip ratio of less than or equal to0.6 (i.e., the radius of the inner, hub end of the trailing edge is nomore than 60% of the radius of the outer, tip end of the trailing edge).

While convention axial blades having high hub-to-tip ratios also requirelarge numbers of blades to extract any significant amount of energy fromthe exhaust, the present blades are capable of extracting a very highpercentage of the dynamic pressure of the high-speed highly tangentialflow entering the turbine wheel. They can do so with a relativelylimited number of blades, thereby limiting the rotational moment ofinertia of the turbine wheel, and therefore providing for fast transientresponse time. Under numerous embodiments of the invention there are 20or fewer blades, and for many of those embodiments there are 16 or fewerblades.

At any given radial location along the blade, the lower and uppersurfaces are each characterized by a camber, and the blade ischaracterized by a median camber, which for the purposes of thisapplication will be defined as a median camber curve 249 extending fromthe leading edge to the trailing edge at a median location equallybetween the upper and lower surfaces, wherein the median location istaken along a lines 251 extending from the upper camber to the lowercamber, normal to the curve 249 along the median camber curve.

The median camber curve 249 comes to a first end at the leading edge245. The direction of the median camber curve at the leading edgedefines a leading-edge direction 253, and is characterized by aleading-edge direction angle β₁ (i.e., a β₁ blade angle) that is theangular offset between the leading-edge direction and a line that isparallel to the axis of rotation and passing through the leading edge(at the same radial location as the median camber), and therefore alsoto the downstream axial component 221 of the supersonic flow. The β₁blade angle is positive when the leading edge turns in to thecircumferential flow component 223 (as depicted in FIG. 5), and zerowhen the leading edge faces directly along the axial flow component 221.The β₁ blade angle can vary over the radial extent of the leading edge.

The median camber curve 249 comes to a second end at the trailing edge247. The direction of the median camber curve at the trailing edgedefines a trailing-edge direction 255, and is characterized by atrailing-edge direction angle β₂ (i.e., a β₂ blade angle) that is theangular offset between the trailing-edge direction and a line that isparallel to the axis of rotation and passing through the trailing edge(at the same radial location as the median camber). The β₂ blade angleis positive when the trailing edge turns in to the circumferential flowcomponent 223 (as depicted in FIG. 5), and zero when the trailing edgefaces directly along the axial flow component 221. The blade angle β₂can vary over the radial extent of the trailing edge.

The sum of the β₁ and β₂ blade angles at a given radial location on ablade defines a turning angle for the blade at that radial location. Theβ₁+β₂ turning angle can vary over the radial extent of the blade.

While the primary scroll efficiently accelerates the exhaust gas streamand thereby provides for a substantial increase in the dynamic pressureof the exhaust gas stream, it does not typically produce a flow with ahigh degree of axial uniformity, as might be seen from a vaned nozzle.The blades of the present embodiment, and particularly the shapes oftheir leading edges, are tailored so that each radial portion of theblade is best adapted to the flow that occurs at its radial location.This type of tailoring is not typical for conventional axial turbines,as they typically have vaned nozzles providing a high level of flowuniformity, and as they have a much higher hub-to-tip ratio that limitspossible variations between the hub and tip flows.

Under the present embodiment, over the majority of the leading edge ofeach blade, the blade angle faces circumferentially upstream withrespect to the axis of rotation (i.e., the β₁ blade angle is positive).Moreover, the β₁ blade angle is greater than or equal to 20 degrees (andpossibly greater than or equal to 30 degrees) at both the hub end of theleading edge and the mid-span of the leading edge (i.e., the leadingedge half way between its hub end and its shroud end). At the shroud endof the leading edge, the β₁ blade angle is greater than or equal to −20degrees (and possibly greater than or equal to −5 degrees).

Additionally, under the present embodiment, over the majority of theradial extent of each blade, the β₁+β₂ turning angle is positive.Moreover, the turning angle is greater than or equal to 45 degrees atthe hub end of each blade. The turning angle is greater than or equal to80 degrees at the mid-span of each blade. At the shroud end of eachblade, the β₁+β₂ turning angle is greater than or equal to 45 degrees.

The chord line 261 (i.e., the line connecting the leading and trailingedge) has a positive angle of attack with respect to the downstreamaxial component 221, i.e., even though the leading-edge direction facescircumferentially upstream with respect to the axis of rotation, thechord line itself is angled circumferentially downstream with respect tothe axis of rotation. In other words, the leading edge iscircumferentially downstream of the trailing edge. This may vary inother embodiments.

The lower surface 241 of the blade of this embodiment is configured tobe concave over substantially the full chord of the blade. Moreover, atthe majority of radial locations, the lower surface is curved such thatit has a range of locations 263 that are circumferentially downstream ofboth the leading edge and the trailing edge.

Static Pressure Drop

A key feature of the present embodiment of the invention is that itprovides the inertial advantages of a typical axial turbine wheel(having a lower rotational moment of inertia than an equivalent radialturbine wheel), while it greatly enhances the ability of the axialturbine to extract the energy of the exhaust gas stream. To accomplishthis, as previously suggested, the present embodiment is provided with avolute that uses conservation of angular momentum to efficientlyaccelerate the exhaust gas stream and convert a significant portion ofthe total pressure in the exhaust gas stream from static pressure todynamic pressure, and further to provide the accelerated exhaust gasstream to an axial turbine wheel at a significant angle.

The turbine blade is configured to extract a significant portion of theenergy of the dynamic pressure from the flow, but not to significantlychange the static pressure of the flow. As a result of the voluteconverting a significant portion of the static pressure to dynamicpressure, and of the wheel extracting most of the dynamic pressurewithout changing the static pressure of the airstream, the turbineextracts a large percentage of the energy in the exhaust gas streamwithout receiving a significant axial load. A typical embodiment of theinvention will be characterized by a static pressure change across theturbine wheel blades of less than ±20% of the static outlet turbinepressure across the turbine for at least some operating conditions ofthe range of standard operating conditions, thereby causing very littleaxial force to be applied to the turbine wheel. More particularly, theturbine is configured to limit the static pressure upstream of the wheelnear the wheel hub to a value that is not greater than 120% of theturbine outlet static pressure, i.e. the pressures differ at the most by20% of the outlet pressure. Some embodiments of the invention arecharacterized by substantially no static pressure drops across therotor, thereby causing only a negligible axial force on the turbinewheel.

Wheel Hub

With reference to FIGS. 5 & 6, the radial size of the turbine wheel hub271 varies along the blade inner hub end 273 from the leading edge 245of each blade 231 to the trailing edge 247 of each blade, and it isuniform around the circumference. More particularly, the hub is radiallylarger at the leading edge than it is at the trailing edge, and the hubis radially larger at an intermediate axial location between the leadingedge and the trailing edge than it is at either the leading edge or thetrailing edge. This increase in thickness forms a smoothly continuoushump 277 that is axially close to the range of locations 263 on theblade lower surface 241 that are circumferentially downstream of boththe leading edge and the trailing edge (i.e., where the median camber isparallel to the axial component of the flow).

The hump 277 is provided in a location in which significant diffusionoccurs, and it prevents the diffusion from exceeding a critical level atwhich flow separation might occur. The potential for this problem isuniquely substantial because of the unique size and shape of the bladesand the high level of kinetic energy of the flow. Because use of thehump helps avoid flow separation, the hump provides for improvedefficiency over a similar wheel that lacks the hump.

Axially Balanced Compressor

With reference to FIG. 2, the compressor housing 207 and compressorwheel 213 form a dual, parallel, radial compressor. More particularly,the compressor wheel has back-to-back oriented impeller blades. A firstset of impeller blades 301 are oriented in a conventional configurationwith an inlet facing axially outward (away from the turbine) to receiveair from that direction. A second set of impeller blades 303 areoriented in a reverse configuration with an inlet facing axially inward(toward the turbine) to receive air brought in tangentially and turnedto travel axially into the second set of impeller blades. The first andsecond set of impeller blades can be manufactured in the form of asingle, integral wheel, e.g., as illustrated, or may comprise anassembly of a plurality of parts.

The compressor housing 207 is configured to direct inlet air to each setof compressor blades in parallel, and to direct the passage ofpressurized gas from each compressor. In this embodiment, the compressorhousing comprises two separate axially positioned air inlets; namely, afirst air inlet passage 305, that is positioned adjacent an end of thecompressor housing to pass inlet air in an axial direction to the firstcompressor blades 301, and a second air inlet passage 307 that isseparate from the first air inlet passage 305. Pressurized air that isprovided by the compressor wheel 213 is directed radially from each setof impeller blades 301 and 303 through a single passage 311 to acompressor volute 313.

This dual-path, parallel, radial compressor configuration, whiletypically being less efficient than a comparable single-path radialcompressor, will operate at higher speeds and might producesubstantially no axial loading in steady state operation. The higheroperating speeds will typically better match the operational speeds ofthe axial turbine.

Synergies

The configuration of the present embodiment is significant for a numberof reasons, and it is particularly effective for overcoming theefficiency limitations that limit the effectiveness of turbochargers onsmall gasoline powered engines, where the practical limitations ofconventional axial turbines render them relatively ineffective forpractical and efficient use.

The present invention provides an effective turbine with large bladesthat can be efficiently manufactured, even in small sizes. Thecomparatively large size and small number of axial turbine blades arewell suited to casting in small sizes when smaller blades might be toosmall for conventional casting techniques. The high speed flow and largeblades do not require manufacturing tolerances that may be limiting whenapplied to a very small turbine.

Singularly, the use of either a no-axial-load turbine or a no-axial loadcompressor is less efficient than their conventional axially loadedcounterpart. Moreover, turbines and compressors are typically configuredto have partially offsetting axial loads. Although these loads are farfrom perfectly matched, they do provide at least some relief from axialloads. If only one component (i.e., either the turbine or thecompressor) creates no axial load, the remaining load from the othercomponent is not partially offset, and even greater axial loads occur,requiring an even larger thrust bearing.

In the present invention, a no-axial-load compressor is combined with ano-axial-load turbine, allowing for the use of much more efficientthrust bearings. It is believed that in some embodiments the thrust loadrequirements may be as small as only 20% of the conventionalcounterparts. Bearings configured to carry such small loads can beadapted to be substantially more energy efficient. As a result, despitethe potentially lower efficiencies of some of the system components, theoverall system efficiency of the turbocharger may be significantlyhigher than in a conventional counterpart.

Other Aspects

While many conventional turbochargers are designed to produce nodownstream swirl, some embodiments of the present invention may beconfigured with blades that produce either negative or even positiveswirl. In designing a turbine under the present invention, theproduction of downstream swirl might be considered of less interest thanin the efficient extraction of energy while producing little or no axialloading.

It is to be understood that the invention comprises apparatus andmethods for designing and producing the inserts, as well as for theturbines and turbochargers themselves. Additionally, the variousembodiments of the invention can incorporate various combinations of thefeatures described above. In short, the above disclosed features can becombined in a wide variety of configurations within the anticipatedscope of the invention.

For example, while the above-described embodiment is configured as aforward-flow turbocharger (i.e., the exhaust gas stream is streamedthrough the turbine wheel so as to come axially out the end of theturbocharger), other embodiments may be configured with a reverse flowin which the exhaust gas stream passes through the turbine wheel in adirection toward the compressor. Such a configuration, while it mightnot fit in the standard spaces allotted for internal combustion engineturbochargers, exposes the bearing housing to less heat and pressure.Also, while the described embodiment uses a wheel with cantilevered(i.e., free-ended) blades that are radially surrounded by an unmovinghousing shroud, other embodiments employing a shrouded wheel (i.e., awheel having an integral shroud that surrounds the blades and rotateswith them) is within the scope of the invention.

While particular forms of the invention have been illustrated anddescribed, it will be apparent that various modifications can be madewithout departing from the spirit and scope of the invention. Thus,although the invention has been described in detail with reference onlyto the preferred embodiments, those having ordinary skill in the artwill appreciate that various modifications can be made without departingfrom the scope of the invention. Accordingly, the invention is notintended to be limited by the above discussion, and is defined withreference to the following claims.

1. A turbocharger configured to receive an exhaust gas stream from anengine configured to operate over a range of standard operatingconditions, and to compress input air into a pressurized air stream,comprising: a housing including a turbine housing; and a rotorconfigured to rotate within the housing along an axis of rotor rotation,the rotor including an axial turbine wheel, a compressor wheel, and ashaft extending along the axis of rotor rotation and connecting theturbine wheel to the compressor wheel; wherein the turbine wheel isconfigured with a hub, and with a plurality of axial turbine bladesconfigured to drive the rotor in rotation around the axis of rotorrotation when the turbocharger receives the exhaust gas stream from theengine, the blades having an axially upstream edge, an axiallydownstream edge, a hub end, and a tip end opposite the hub end, thetrailing edge being characterized by a radius at the hub end and aradius of the tip end; wherein the compressor wheel is configured tocompress the input air into the pressurized air stream when the rotor isdriven in rotation around the axis of rotor rotation by the turbinewheel; wherein the turbine housing forms an inwardly spiraling turbineprimary-scroll passageway that turns in an axial direction; and whereinthe radius at the hub end of the turbine wheel trailing edge is no morethan 60% of the radius of the tip end of the turbine wheel trailingedge.
 2. The turbocharger of claim 1, wherein the turbine blades areeach characterized by a blade turning angle at the hub that is greaterthan or equal to 45 degrees.
 3. The turbocharger of claim 2, wherein theturbine blades are each characterized by a leading edge blade angle atthe hub that is greater than or equal to 20 degrees.
 4. The turbochargerof claim 3, wherein the turbine blades are each characterized by aleading edge blade angle at the hub that is greater than or equal to 30degrees.
 5. The turbocharger of claim 3, wherein: the turbine blades areeach characterized by a blade turning angle at an intermediate radiusbetween the hub and the tip that is greater than or equal to 80 degrees;and the turbine blades are each characterized by a leading edge bladeangle at the intermediate radius that is greater than or equal to 20degrees.
 6. The turbocharger of claim 5, wherein the turbine blades areeach characterized by a leading edge blade angle at the hub that isgreater than or equal to 30 degrees, and by a leading edge blade angleat the intermediate radius that is greater than or equal to 30 degrees.7. The turbocharger of claim 2, wherein the turbine blades are eachcharacterized by a blade turning angle at an intermediate radius betweenthe hub and the tip that is greater than or equal to 80 degrees.
 8. Theturbocharger of claim 7, wherein the turbine blades are eachcharacterized by a blade turning angle at the tip that is greater thanor equal to 45 degrees.
 9. The turbocharger of claim 2, wherein theturbine blades are each characterized by a blade turning angle at thetip that is greater than or equal to 45 degrees.
 10. The turbocharger ofclaim 1, wherein the turbine blades are each characterized by a bladeturning angle at an intermediate radius between the hub and the tip thatis greater than or equal to 80 degrees.
 11. The turbocharger of claim10, wherein the turbine blades are each characterized by a leading edgeblade angle at the intermediate radius that is greater than or equal to20 degrees.
 12. The turbocharger of claim 10, wherein the turbine bladesare each characterized by a blade turning angle at the tip that isgreater than or equal to 45 degrees.
 13. The turbocharger of claim 1,wherein the turbine blades are each characterized by a blade turningangle at the tip that is greater than or equal to 45 degrees.
 14. Theturbocharger of claim 13, wherein the turbine blades are eachcharacterized by a leading edge blade angle at the tip that is greaterthan or equal to −20 degrees.
 15. The turbocharger of claim 1, whereinthe turbine wheel is limited to no more than 20 blades.
 16. Theturbocharger of claim 15, wherein the turbine wheel is limited to nomore than 16 blades.
 17. The turbocharger of claim 1, wherein theturbine wheel blades are cantilevered from the turbine wheel hub. 18.The turbocharger of claim 1, wherein the turbine is configured to limitthe static pressure upstream of the wheel near the wheel hub to a valuethat is not greater than 120% of the turbine outlet static pressure forat least some operating conditions of the range of standard operatingconditions.
 19. A turbocharged internal combustion engine system,comprising: an engine configured to receive a pressurized air stream andto produce an exhaust gas stream, the engine being configured to operateover the range of standard operating conditions; and the turbocharger ofclaim 1, the turbocharger being configured to receive the exhaust gasstream from the engine when operating in the standard operatingconditions, and to compress input air into the pressurized air streamreceived by the engine.
 20. The turbocharged internal combustion enginesystem of claim 19, wherein the inwardly spiraling primary-scrollpassageway substantially forms a convergent passageway that turnsaxially downstream and spirals inward enough to cause the input air toachieve supersonic speeds when reaching the upstream edges of theturbine wheel blades for at least some operating conditions of the rangeof standard operating conditions.
 21. A turbocharger turbine wheel foruse in a turbocharger configured to receive an exhaust gas stream froman engine configured to operate over a range of standard operatingconditions, and to compress input air into a pressurized air stream,comprising: a hub; and a plurality of axial turbine blades configured todrive the rotor in rotation around the axis of rotor rotation when theturbocharger receives the exhaust gas stream from the engine, the bladeshaving an axially upstream edge, an axially downstream edge, a hub end,and a tip end opposite the hub end, the trailing edge beingcharacterized by a radius at the hub end and a radius of the tip end;wherein the radius at the hub end of the turbine wheel trailing edge isno more than 60% of the radius of the tip end of the turbine wheeltrailing edge.
 22. The turbocharger turbine wheel of claim 21, whereinthe turbine blades are each characterized by a blade turning angle atthe hub that is greater than or equal to 45 degrees.
 23. Theturbocharger turbine wheel of claim 22, wherein the turbine blades areeach characterized by a leading edge blade angle at the hub that isgreater than or equal to 20 degrees.
 24. The turbocharger turbine wheelof claim 22, wherein the turbine blades are each characterized by ablade turning angle at an intermediate radius between the hub and thetip that is greater than or equal to 80 degrees.
 25. The turbochargerturbine wheel of claim 21, wherein the turbine blades are eachcharacterized by a blade turning angle at an intermediate radius betweenthe hub and the tip that is greater than or equal to 80 degrees.
 26. Theturbocharger turbine wheel of claim 25, wherein the turbine blades areeach characterized by a leading edge blade angle at the intermediateradius that is greater than or equal to 20 degrees.
 27. The turbochargerturbine wheel of claim 21, wherein the turbine wheel is limited to nomore than 20 blades.
 28. The turbocharger turbine wheel of claim 27,wherein the turbine wheel is limited to no more than 16 blades.
 29. Theturbocharger turbine wheel of claim 21, wherein the turbine wheel bladesare cantilevered from the turbine wheel hub.